Sonic generators used to convert electrical energy into acoustic and kinetic energy for transmission to fluid mediums are known. Such devices are illustrated, for example, in U.K. Patent Specification 2,152,728 to Bodine and U.S. Pat. No. 2,468,515 to Robinson.
The sonic generators there disclosed, however, suffer disadvantages. For example, the apparatus taught by Bodine in the aforementioned U. K. patent utilizes resonant drives similar to the types described in Bodine's U.S. Pat. Nos. 3,633,877, 3,684,037, 3,360,056 and 4,265,129. Such drives limit the upper range of frequencies such units can transmit to a fluid medium. Since the transmission of acoustic power between the resonant bar and the fluid is less efficient at lower frequencies, the frequency limitation lowers the acoustic efficiency of the Bodine apparatuses. Bodine also necessarily utilizes a coupling between his orbiting mass oscillators and his resonant bar which attempts to isolate the inertial forces of the vibrating bar from the oscillators. Without such a coupling, the magnitude of the forces generated are sufficient to cause relatively rapid failure of the orbiting mass oscillator drive motors, be they hydraulic or electric.
The above-identified Robinson reference teaches restraining a resonating bar with steel bushings at respectively oppositely located end portions of the bar. Such supports cause energy to be lost through the steel bushing support structure which energy would be better utilized in the fluid medium. This is so since Robinson's method does not allow for free and unrestrained vibration of the resonant member. Rather the resonant member is forced to assume a specific mode shape by the steel retaining bushings. Robinson's structure also imposes a very high stress concentration on the resonant member both at the support point and at the point of maximum bending stress. Such stress concentration can eventually cause unnecessary damage to the resonant member and/or premature failure. It also limits the mechanical stresses that can be sustained by the member without failure.
Existing vibratory grinders generally consist of a rigid housing containing the grinding media such as steel balls mounted on a spring system for support and vibration isolation. The vibration is transmitted through a rotating unbalanced shaft that is rigidly attached to the body of the grinding unit or through an electric motor using an integral unbalanced weight mounted directly to the housing.
The unbalanced shaft embodiment is typically driven by a standard electric motor through a cardan shaft to isolate the motor from the vibration or the eccentric weight which is integral to the electric motor. If the latter, the electric motor vibrates with essentially the same intensity at the grinder itself.
When inducing vibration with an eccentric weight, forces are introduced which must be transmitted from the rotating member through bearings to the body of the grinder. These forces increase with the square of the vibration frequency which practically limits the frequency of the vibration to 30 Hz in commercial scale applications because of the very large force required to vibrate the mass of the grinder unit as the frequency increases. As the rotational frequency increases, the roller bearing becomes limited in load carrying capacity.
When existing vibratory designs are referred to as resonant grinders, it merely means that the vibration frequency corresponds to the natural frequency of the spring-mass system used to isolate the grinder. Resonant operation of such grinders will increase the energy efficiency in the system. However, it is difficult to obtain the necessary vibration amplitude for the grinder with the required spring stiffness for the large grinding masses.